Vacuum control of clutch capacity



May 12, 1942. J. DoLzA VACUUM CONTROL OF CLUTCH CAPACITY Filed 0G12. 18, 1957 '7 Sheets-Sheet l ///V///////2 Munn, QN F l,

May 12 1942 J. DoLzA y 2,282,949

VACUUM CONTROL OF CLUTCH CAPACITY Filed 001'.. 18, 1957 7 Sheets-Sheet 2 May `12, 1942. J. DoLzA 'VACUUM CONTROL OF CLUTCH CAEACITY '7 Sheets-Sheet 5 Filed 001,. 18, 1937 May'=12, 1942. J. DO LzA 2,282,949

VACUUM CONTROL OF CLUTCH CAPACITY Filed oct. v18, 1937 7 sheets-sheet. 4

fahr? @ggf Maylrz, 1942. J, DOLZ `2,282,949

VACUUM CONTROL. oF CLTCH CAPACITY Filed Oct. 18, 1957 7 Sheeks-Sheec 6 gmc/who@ May-12, 1942.

. liled oct. 18',V 193'? 7 Sheets-Sheerl '7 ENG/NE Sp-ED ERM Patented May 12, 1942 VACUUM conraoL or cmrcn caracrrr .form no1, runt, Mich., magno;- fo Generar nrotors Corporation, of Delaware Detroit, Mich., a corporation Application October 18, 1937, Serial No. 169,535 14 Claims. (Cl. 192-.01)

The invention relates to the control of variable speed ratio in transmissions driven by engines and connected to a varying load, more particularly to present day motor vehicles which are driven by internal combustion engines, and Aare equipped with step ratio gears requiring transfer of torque-from one of a group of -driving paths toanother. v It relates further to mechanism for automatically selecting speed ratio, and more particularly to forms of transmissions in which the transition interval occurs without i'ull release of i torque, or with torque overlap.

1.. An objectof my invention is the establishing f@ shift sequence in which exceptionally smooth transfer of torque is provided by the coordination of fluid pressure ratio shifting means withregulatory means subject to a measure of torque or to torque demand.

A; further object is the providing of regulatory means instantly available and operative ,at the will of the operator to modify the establishing of drive in a given shift interval according to the existing degree of engine torque.

An additional object is the provision ofmeans subject to the degree of engine intake manifold vacuum applied to the controls regulating the engagement of drive during the shift interval, whereby predetermined vforces acting on fluid pressure valving are correlated with the driving conditions.

Other objects and advantages to be derived from the use of the :invention hereindisclosed reside in the interrelation and methods of operation of the parts described, and will become apparent upon inspection of the following speciiication when read with reference to the accompanying drawings wherein the preferred embodiments of the invention are illustrated.

It is expressly understood, however, that the drawings are for the purpose of illustration only, and are not to be taken as a definition of the limits of the invention, reference hereunder being noted for this purpose to the appended claims. In the drawings:

Figure 1 is a stereographic sectioned projection of a transmission installation in a motor vehicle, showing the sequence arrangement of units from engine to load shaft diagonally from left to right.

Figure 2 is an enlarged view of the forward unit of Figure 1, shown in sectioned projection.

Figure 3 is a similar enlarged view of the rear unit of Figure 1.

Figure 4' is a schematic view of the control and actuating mechanism of Figures l, 2 and 3, providing detail of the valving controls and associated parts.

Figure 5 is an enlarged view of the servo and lubrication pump'arrangement of Figure 1` Figure 6 is an elevation section view of the mounting for the .valve controls ,ofv Figure 4, and shows Athe `external connections linking wi the driver and governor connected control members of Figure 1.

Figure 'I shows the driver selection means attached to the steering column oi' the vehicle, in

projection view.

Figure 8 is a modication of the vacuum responsive means of Figure 4 used to vary the rate and degree. of loading pressure on the direct drive clutches in units B and C.

Figure 9 is a chart illustrating in heavy 4line .the form of the control characteristics provided by the vacuum response of the mechanism of Figures 4` and 8.

In Figure 1 the projected schematic sectioned view of the transmission assembly of my example discloses a first variable speed unit adjacent the engineconnected main clutch shaft 5, having output shaft 8 carrying splined gear I9 engageable with reverse idler gear I8, or with 'jaws 1' of gear 1 integral with shaft 5, and

constantly meshed with countershaft gear I8 rotatable with gear I1 driving reverse idler gear I8. Countershaft body 2l rotates a rst gear |13 of the servo pump drive assembly,hereinafter designated by letter P, and shaft 8 drives a ysecond gear-|10 of the servo pump drive assembly, to be described later. Body 20 rotates on I5.

Gear 9 is shifted axially by fork |00 attached to slider |04 of rod |0I, and rocker |02 moving in slide |05, the shaft |03 of rocker |02 projecting externally from the casing 2. The above head gearset assembly constitutes a forwardneutral-reverse shifting unit,` hereinafter designated by letter A.

carries clutch plate 33 mating with plate 36.

held to rotate 'with drum 28 and presser plate 14. Springs 88 tend to disengage plate 33-36,

' and pistons 12 sliding in cylindrical spaces 1| -in drum portion 29 are arranged to press centering spindle 13 and presser plate 1I to overcome springs 88 and build up capacity in the clutch formed by plates 33-36. Fluid pressure is fed through external pipe 218', gland passage 281 and drilled passage 19 to load plates 33-36 to sustain direct drive in the unit constituted by the described planet gearing assembly, hereinafter designated by letter B.

The piston pins 13 transmit loading effort to presser plate 1I, plates 36 and 33, the reaction pivoted thereto at 98a, acted on by pivoted to the casing at |94.

being supported by the inturned flange of gear 25 against rotation, whereby reduction gear drive through the gearing is made possible.

When clutch 33-36 is not driving, brake band 88 supported in the transmission casing as shown in Figure 1, is applied, holding drum 28-29, sleeve 26 and reaction sun gear 25 from rotation. With power applied from the engine to shaft 8, and annulus gear |2, the planet gears 24 rotate on their spindles, and cause planet carrier 22 to rotate at a reduced speed, imparted to shaft 2|. Shaft 2| extends through unit B, to serve as the power input member for unit C shown in Figure 3 and having two driving sun gears 31--38 aiilxed thereto. Forward planet gears 44 meshing with sun gear 38 are rotatably mounted in carrier 54 whose drum extension 52 terminates in internally toothed annulus 5|. Annulus gear 42 meshing with planets 44, is integral with plate '39 fixed to drum 39. 'Ihe final output shaft 58 one of two positions, for `admitting servo line pressure from servo main 238 and port 266 to port 261 and line 218 for actuating unit B for direct drive; or alternately for 'cutting off servo main 238 and releasing fluid pressure from line 218 to exhaust port 268, for establishing drive through the gearing.

Likewise valve |68 occupies one of two positions for controlling transmission unit C; the first position connecting servo line 213 and port 263 to port -262 and line 219 for establishing direct drive; and alternately cutting off line 213 while opening pipe 219 to exhaust through port 258 for establishing drive through the gearing.

is integral with planet carrier 45, rotatably supporting planet gears 43 meshing with sun gear 31 and annulus 5|. Clutchy hub 59 splined to shaft 2| is splined externally to support clutch plate 68 mating with plate 55 carried by studs in web 39' and end wall 56 of drum 39, releasing springs 88 similar to springs 98 of Figure 2 serving the same function as springs 8,8 of unit B, named in this specification `as the front unit. Pistons 16 may press on presser plate 18 mounted to rotate with drum 39, through rods 11, the pistons occupying cylinders 15 cut in drum web 56, ailixed to section 39 and web 39. Actuation of brake 98 of unit C is through thrust rod |98 rocker |93 The clutches in the drawings herewith are shown schematically as of single plate type, but the invention is applicable by those skilled in the art to multiple plate constructions such as described in S. N. 45,184, filed Gctober 16, 1935, to E. A. Thompson, now matured as U. S. 2,193,304, issued March 12, 1940.

The invention herewith is directed to controls for transmission clutches and not to clutches per se.

F'luid pressure is fed to pipe 219 to gland 289, and through passages 281 and 19' to cylinders 15, for loading plates 55-68, to establish direct drive in unit C, known in this specification as the rear unit.

Brake 98 as shown in Figure 3 encircles drum 39, and may lock it, and reaction annulus gear 42 against rotation, when reduction drive in unit C is desired. at which time clutch 55-68 is released. n

Alternate operation of the actuating elements for the forward speed ratio drive of units B and C is obtained from the following tabular pattern:

Unit B Unit C Ru l lo Clutch Brake Brake Pump assembly P, shown in Figure 1, operates continuously as long as either of shafts 5 or 8 have rotation, as will be understood by inspection of Figure 5, which shows the parts in transverse section. The main outletfrom the pump is controlled by regulator valve 288which also controls lubrication pressure to the force feed oiling system, not essential to the invention of this specification. Servo main 238 leads from the outlet pressure space of the pump assembly and'from valve 288 to ports 266 and 263 of valves |58 and |68 respectively, and also through lin'e 338' to port 265 of compensator valve 328 of Figure 4.

For convenience in assembly and manufacture, valves |58, |68 and 328 are mounted in a common valve body 218 adjacent to or integral with transmission casing 2.

As shown in Figure 4, the shift pattern of valves |58 and |68 with respect to the aforementioned shift-actuation table, is as follows:

Valve Valve Down.

It will be further understood that when theboss of the valve is raised beyond port 283 leading to check valve 2|9 and servo main 238. At this time'leakage through slots 288-2|4 per'rnits flow of oil to port 284 and lubrication line 228. .With a further increase in pressure the lower boss will open port 284 to full pressure, at which time the upper boss will open port 285 and line |91 to set up a balancing pressure on the upper face of the lower boss, whereafter a pressure balance between adjustable spring 28| and the pump pressure is maintained at a fairly constant pressure point, in that overpressure will blow off the surplus through exhaust port 286, and relief valve 2|1. The overpressure relief occurs when the lower edge of the upper boss permits ow to port 2|1. I'he primary purpose of this valve is to afford regulation to pump line servo pressure,

and the secondary purpose is to move rapidly to open servo line position with the first incremental rotation of the pump. Other forms of regulating valves may be used without departing from the principles of my invention, the requirement herein being that uniform servo line pressure be sustained in servo main 238. Screw 2|5 adjusts tension of spring 20|.

As shown in Figure 6, valve |68 for unit B is positioned by pivoted bellcrank |60, movable clockwisefor shifting the valve to up position, and counterclockwise for shifting it to down position. Pin |59 of member |60 intersects slot |55 of camplate |26 pivoted adjacent to the pivot |58 of member |60, so that rocking of |26 counterclockwise may cause valve |68 Ato move to the down position, according to the radial distances of points on cam slot |55 from the 'center of shaft |52 on which |26 may turn. External lever |5| of shaft |52 is connected to manual shifter elements 3|0, |08, 308, 306, 305 and hand lever 30| movable over speed ratio sector positions as indicated on the sector plate of Figure '7. Line 220 of Figures l-and 5 transmits lubricant to the transmission unit gears and bearings. Leakage passes 208 and 2|4 provide initial flow of oil in the motion of valve 200. Inlet |91 feeds to port 205. Pipe |9| is the suction inlet for pump P, leading to suction space |10. Idler pump gears |82 and |84 mesh with rotor gears |1| and |80. Sleeve |18 receives drivefrom gear meshing with gear |13. Check valve 2|8 prevents teo rapid flow of lubricating oil in line 238.

'I'he inner end of shaft |52 carries aflixed lever 4|5 pressing on spring 4|6, which in turn may press on piece 4|1 pivoted to plate |26, and thereby transmit rotational force through spring 4|6 to camplate |26, rocker arm |60 and valve |68. Stop pins on plate |26 prevent departure of levers 4|5 and 4|1 from the margins of camplate |26.

The hand control lever 30| may therefore shift valve |68 to a direct or to a gear drive positionl for the rear unit C.

In Figure 4 valve |50 receives servo pump pressure from line 238 at port 266. Valve |50 is toggle operated as shown, through toggle arms 13S-|39 pivoted at |31, held by spring |48, and operated by rod'l I3, but biased by spring Il toward the active right-hand position. With the valve positioned as shown in dashed lines in Figure 4, pump pressure from 266 may be applied through port 261 to line 218, to actuate the elements of the brake cylinder 282 and clutch parts connected to passage 19. In the alternative full line position, valve |50 connects port 261 to the exhaust port 268, and servo port 266 is cut off.

Valve |50 is moved by automatic and manual means, as shown in Figure 6.

Automatic shifter rod 3 is pivoted at its forward end to equalizer bar pivoted to governor rod ||0 at ||2. Idler lever |35 pivoted at |35 to valve body 210 is pivoted to rod I0 to limit the motion. Extension 420 of |35 is arranged to intersect the movement of lever 4| 1 pivoted on camplate |26.

Cross-shaft 24| of Figures 1 and 5 is driven by gear meshing with gear |14 of shaft 8. Governor weights swing outward with rotation of ange 244 of shaft 24|, the weightarms 260 camming sleeve 233 connected to external fork 350 against the action of governor springs 252- 254, thereby rotating shaft attached to the fork, and rocking lever 352 attached to the shaft and pivoted to rod ||0 at 353. The governor assembly will hereinafter be noted by letter G.

It will be seen that through the described linkage, as in Figure 1, the variations in speed of shaft 8, connected to the vehicle engine when the drive is forward, will be transmitted to equalizer bar The accelerator pedal 303 for the engine is connected to lever |32 anixed to shaft |20 1I of Figure 6 mounted to turn in the casing 2, through rod 355, lever358, shaft 358, lever 364, and rod 36|. The effective length of rod 36| may be adjustable by any commonly known means, as required for effective operation. The inner portion of shaft |20 carries aillxed lever |3|. Lost motion lever ||6 is pivoted to the casing, and arranged to swing in a path intersecting the motion of lever |3|. Lever ||6 is drilled at its pivot point 4|| and connecting passage 4|2 leads to cylindrical space 4|4, in which slides projecting piston 4|3 engaging the swinging end of lever I3 Pressure line 4|0 joins pivot 4|| to pressure lead 218 of the rear unit, as in Figure 4, so that whenever the rear unit is in direct drive, piston 4|3 couples to lever |3| by the volume of fluid existing behind the piston in cylinder 4|4. Pin 5 of lever |'I6 may engage equalizer bar at notch I4.-

Movement of the pedal 303 for advancing the engine throttle in the normal way through lever 358' and throttle rod 363, will at the same time exert a force through the linkage just described, upon pin 5, tending to cancel or oppose the effect of governor force from rod I0, by shifting the fulcrum point farther to the left, thereby requiring a higher governor speed for a permitted shift of valve |50 to direct drive position for the front unit B. 'I'he throttle pedal connected ratio control linkage above described will hereinafter be designated by letter T.

At the lower portion of Figure 2 is a schematic view of the assembly of control and servo de-` vices constituting the exerrplary system for my invention in which brake cylinder 282 is a housing for brake piston 28| sldable on rod 260, piston 285 attached to rod 280, springs 01, 81a, 81h; abutment 300 xed to retainer 283, and sliding abutment 286 and 288. This assembly is for operating the brake of the front unit B.

The thrust of springs 81, 81a is exerted on piston 28|, and through a shoulder of rod 280, on rocker 393 pivoted to the casing 2 at 394, whereby notch 382 of rocker 393 and thrust rod 390 pivoted to brake member f80 at 80a may apply brake 80 to drum 28 of the front unit. Pressure of l line 218 may relieve the spring force on brake 80 by moving piston 28| to the right in the cylinder 282, while at the same time is acting through pipe 218 to load clutch plates 33 and 36 through presser plate 14 and pistons 16 of the front unit B.

The spring forces against which the servo line pressure from 218 is required to work are of predetermined magnitude, so that the resistances felt by the pump line pressure in acting against the springs provide a scalar change in the degree of pressure exerted on the clutch plates 33 36 through pistons 12 connected to line 19 and also to 218. It will be seen that if means are provided to apply pressure in line 211 to piston 285, the resistance of the spring system to pressure from line 218 acting on piston 28| will be lessened and varied with the variation of pressure in pipe 211.

Figures 2 and 4 show the construction of the plunger 323 operated by movement of the engine throttle pedal sus 0f Figure 1 through changes induced in the degree of engine vacuum. Differential valve 320 slides in bore 3| 8 of valve body 210, the external shell of plunger 323 being bored out internally to t collar washer 321 sldable on adjacent end of stem 32| of valve 320. Lock ring' 340 prevents washer 3,21 from further motion induced by tension in'spring 322.

The uppermost lead 3|8, as in Figure 8, connects thehead of stem 3|4 to pressure line 213 of the front unit. The ported passage 265 connects to line 239 (213) receiving net pump output pressure.r Ported space 335 is joined by line 216- to compensator lines 211 and 211', and cross-connected to space 338' by passage 331. Spring 322 reacts between the lower face of valve 320 and the recessed portion of plunger sleeve 323, guided by stem 32| of the valve.

With no pump pressure available in either servo line 218 or 219, brakes 80 and 90 are active to establish low speed gear drive in both of the front and rear units B and C by virtue of. brake loading springs 81, 81a, 81h, and 91, 91a, 91b.

When the hand control 30| is shifted to move valve |68 to the dashed line position of Figure 4, ports 262 and 263 are no longer connected, and pressure inlet port 263 is shut 01T, draining line pressure from cylinder 292 controlling the brake 90 of the rear unit, through exhaust port 260. Pipe 4|0 joins pipe 218 to passage 4|2 (Fig. 6).

The compensatorpressure inlet 265 is opened to space 335 and to line 216 feeding both compensator lines 211' and 211 connected to act on compensator pistons 285 and 294 of Figures 2 and 3 of the front and rear units, respectively. Balanced pressure on both ends of boss 339 of valve 320 is permitted by passage 331 and ported space 338.

Sleeve 323 may slide in bore 3|9 of valve body 210 and pick up retainer collar 321 to vary the stress of spring 322, whereby the effective aperture between the upper face of boss 339 of valve 320 in space 335 is varied.

Closure of the opening between boss 339 and the seat 355 in space 335 diminishes pressure in line 216 and in connected compensator lines 211 and 211, while conversely, opening of the valve 320 allows the servo pump pressure to be exerted in both of the compensator lines. Exhaust port 336 permits oil to escape from space 355 and line 216.

The hand control of Figures 1 and 7 embodies these members; lever 30|, rod 305, lever 306, clevis 301, rod 308, clevis 309, rod 3| 0 and lever |5| which may rock valve |68 through the mechanism of Figure 6. Indicator sector plate 302 shows the proper shift positions of lever 30|.

'I'he distance through which sleeve 323 may move to effect the compensation pressure inlines 211 and 211' is determined by diaphragm 325 of Figures 4 and 8 and spring 3|9 attached to sleeve 323, mounted in casing 326 and held therein by clamped shell 328, the casing being -attached to valve body 210 by screwed flange 329. Nipple 33| in diaphragm shell 328 is connected to the engine intake manifold 330 through pipe 332 and check valve 333, so as to transmit an index of the engine torque conditions to the sleeve 323. Ring 326 is attached to body 210 and flange 329.

When the engine is stopped, there is no oil pressure acting on the valving, and no vacuum effective on diaphragm 325. The starting of the engine creates a vacuum in line 332 and inside the shell 328, drawing diaphragm 325 down, against the action of spring 3|9 and increasing the compensator port opening of valve 320 between ports 265and 335. As oil pressure from pump P becomes available inline 265, the presclutch 55-60, the compensating pressure behind piston 294 due to the permitted full flow of compensation pressure by valve 320, allows the clutch pressure in line 19 to act on clutch discs 55 and 60 at a lower value than if the opposition of brake springs 91, 91a, 9117, and 91e were fully effective.

When control valve |68 for the rear unit is Line 219' transmits servo'line pressure from cylinder 292 to clutch cylinders 15 from line 219,

by way of gland passage 281 and drilling 19'.

Spring 284 absorbs the first increment of motion of compensator piston 285. Strap 293 is a retainer for the springs 91.

If. however, the engine `throttle be opened before the valve |68 is moved, the reduction in engine vacuum in manifold 330, allows spring 3|9 between shell 328 and-diaphragm 325 to shift valve 320 toward closing of port 216 with respect to line 265, which reduces the compensation pressure in lines 211 and 211', thereby establishing a. much higher pressure value acting on clutch pistons 18 and discs 55 and 60, through the direct resistance action of the springs 81 or 91, on the head of the fluid column of the pump. Abutment 214 transmits primary thrust of pin 291 to 91o.

The capacity of the clutch 55-60 is controlled by the degree of engaging pressure which sustains it. When, for example, the line pressure of the pump P is divided between pistons 29| and 294,-and springs 91, 91a, 91b are not fully effective to establish full pump pressure on clutch pistons-16, the capacity of clutch 55-60 may have a torque value of :c foot pounds, at which time the full range of compensation effect may be present.

Now if compensation pressure be cut olf from piston 294, the full capacity of clutch 55-60 is obtained ina very short time interval, building up to a torque value greater than x, or .'c-i-v. Minute changes in between these pressure and capacity levels are therefore predetermined by variations in the degree of engine vacuum, as created by the drivers opening and closing of the engine throttle, and by the driving condi-` tions which likewise affect the variation of engine vacuum.

Likewise. referring back to Figure 2, compensator pressure in line 211' is exerted on compensator piston 285 of the servo actuator for the front unit.

Conversely, when the car driver advances the engine throttle, the degree of vacuum in manifold 330, pipe 332, and shell 326 diminishes, whereupon spring 3|9 tends to shift diaphragm 325 and valve 320 to closed position, wherein the compensation pressures being no longer sustained, the full pump pressure may build up in lines 218 and 219, according to which of valves |50 or |68 is moved to its direct drive compelling position.

The gradual change of degree of vacuum under these circumstances may be applied to graduation of the opening of compensator valve 320, in a series of infinitely small steps, which translated into pressure effect upon clutches 33-36 and 55-60, provides an exceedingly smooth engaging action, so that there are no discernible lurches or jerks in the engagement of either clutch involved in a speed ratio shift sequence. in either ascending or descending ratio changes.

In Figure 8 the compensator valve 320 of Fig--v ure 4 is moved by spring 322, sleeve 323, andv arm |22 of shaft |23 mounted in valve body 210. Shaft |23 projects externally from body 210, and

assestato is rocked by attached lever |2| pivoted to diaphragm rod 400 attached to diaphragm 325 mounted in casing 328' and shell 328', whose extension carries nipple 33| and affords a seat for spring 3|9' bearing against the diaphragm. Rod 400 is a loose fit at port 432 in shell casing 326 so that if desired, the release or intake of air in the compartment between diaphragm 32.5' and casing 326' may be regulated to dampen or to increase the sensitivity of the action of the vacuum from manifold vacuum lead 332 on the compensator valve 320.

The center of shaft |23 is so taken with respect to the center of valve 320 and plunger 323, that rapid initial motion followed by slower motion in the stroke of plunger 323 with respect to variation of vacuum upon diaphragm 325', may be obtained.

If converse action is required, the center of |23 is located above the point of contact of |22 with 323. This expedient makes it possible to extend the use of my invention to a wide range of power plant installations wherein torque responsive control conditions with respect to available power and load to be driven are Widely different from those experienced in customary passenger car practise.l

The purpose of this control is to proportion the rate of clutching engagement or capacity change to the degree of torque demand existing at the interval when the clutch is selected to drive.

If the torque demand is falling off during the clutch engagement interval, the full pressurev available from the servo pump might become effective before the relative rotations and inertias of the parts had diminished in accordance, whereupon the clutch connected elements would be joined quickly at full loading capacity. This would administer a shock to the transmission system, and cause a lurch of the vehicle, since the absorption period for momentary differences in torque at the clutch discs is very short.

At this point, the vacuum responsive valve 320 interposes the diminishing torque demand factor, and the rise in engine vacuum moves diahragm 325 to compel compensator valve 320 to shift toward an opening position, wherein increasing compensation values operating on compensating pistons 285 and 29d are obtained.

The rst increment of accelerator pedal motion from idling causes the degree of vacuum in manifold 330, pipe 332 and shell 328 to decrease, permitting spring 3|9 to load light spring 322, so as to oppose the force of fluid pressure acting on the upper face of valve 320. This results in a graduating of the orifice between the lower lip of port 335 and the upper face of valve 320, restricting the pressure ow from space 334 to outlets 2l6-2ll, available to create pressure on compensating piston 294 in the cylinder space between abutment wall 295 in cylinder 292.

At full accelerator pedal position the existing degree of vacuum is opposed by spring 3|'9 to an extent such that the end of stem 32| meets the inner end wall of plunger 323, and positive closure of the iiow from space 338 to piston 294 may occur. n

The chart of Figure 9 provides a series of characteristic internal combustion engine torque curves for varying conditionsof engine speed, throttle opening and degree of engine intake manifold vacuum.

The dashed lines represent the engine torques at different engine speeds at given throttle openings, as follows:

Curves Throttle a-a Partial. b b Quarter. c-c Half. d-d; Three-quarters `e. Full.

It will be noted that these curves fall olf rather sharply with increased speed.

The fullV line curves represent the engine torques at different engine speeds at given degrees of engine intake manifold vacuum, diminishing with torque from curves I to IV. These curves are relatively fiat and yield a close approximation to the `torque Values.

'I'his set of curves demonstrates the utility of my invention in providing a uniformly varying change of degree of vacuum with torque, from which resultant force, the coordinate applications of Figures 4 and 8 are obtained. 'Ihe close agreement of the vacuum values withthe engine torque makes the application of my invention a practical utility, and yields an instant response when operating over highways in territories of considerable gradient, so that the ratio shifting clutches are provided, with a torque capacity in close accordance with the actual need.

The arrangement of gearing, clutching and braking shown herewith, provides a. ratio shift sequence in which during forward drive, the

transitions from one ratio to another always occur with a given mean torque value between input and output. In other words, at no time in forward speed drive is there a neutral, or no-drive condition while ratio is being shifted. The overlapping of torque among the shifter elements such as the clutches and brakes used in the illustration provides an extremely smooth method of passing through the transmission shift intervals, but practice has disclosed that supplementary means to adjust the torque capacity of the clutches about the take up drive, in accordance with the existing torque demand is required in order to absorb the existing inertias of the rotating parts, and to avoid abuse of the friction members which have to carry the torque.

It has been shown in the clutch art that regulation of the rate of clutching engagement by varying the position of the engine throttle provides a certain control over clutching action, but this, however, has been applied to friction clutches pre-biasedfor driving, the control being upon the permitted rate of engagement. In my invention the ratio shifting clutches are pre-biased for disengagement, and I superimpose my method upon a system in which automatic and manual controls not only initiate the clutching action but also apply a given rate of pressure build-up to lestablish a predetermined torque capacity level for the clutch or clutches involved.

The end point of the clutching action on thephase of the clutch control cycle 1 derived from predetermined throttle opening and driving conditions against a spring, for example, item 3|9 of Figure 4, of given rate, and likewise against a predetermined the smoothness ofr fluid pressure force act. f

ing in opposition upon valve 320, so that for the range of movement of the compensating control, an extremely Afine adjustment of the fluid pressures acting on the clutches 33-36 and 55--60 is obtained. 'I'his method is in no way sensitive to wear of the clutches, in that pressure of the constantly operating pump is always available to take up slack due to wear, by creating a volume increment of corresponding value.

Variations in viscosity of the fluidl used in the system can only affect the friction of valve 320 in bore 3|9', the fluctuations of viscosity being of magnitudes of less effect than the calculated forces available from vacuum line 332, diaphragm 325, spring 3l9, spring 322, and pressure lines 265 and 3 I 8.

Among the structures to which the present invention is applicable are those described and shown in the publication known in the automotive trade as Automotive Industries, for May 29, 1937, on pages 806 to 809 and 823.

Thus I have provided a novel driving mechanism and control which is effective to achieve the objects above enumerated, which possesses advantages in manufacture l and servicing, and which is adaptable to a wide variety of use and applications. For example the system shown is useful for driving rail cars, aircraft propellers, superchargers for induced air combustion draft, tractors, farm and excavating machinery. Wherever a drive between a variable speed and power engine and a variable load is required the demonstration of my invention is applicable.

It is herewith acknowledged that changes from my construction and arrangement of parts will suggest themselves to those skilled in the art, but it is understood that such means are within the scope of the invention herein disclosed and as defined in the appended claims.

I claim:

1. In automatic variable speed mechanisms, in combination, an engine. a speed control for said engine, multiple variable speed gearing driven by said engine and connected to a load, multiple friction devices arranged to establish different driving speed ratios within said gearing including direct driving coupling clutches, fluid presl an engine, a speed control for said engine, multiplevariable speed gearing driven by said engine actuable by fluid pressure means, a valve controlling one unit of said gearing movable into two positions, one for admitting fluid pressure to, and the other for releasing fluid pressure from said means, a second valve controlling by-passed fluid from said flrst named valve, and a device responsive to engine vacuum interposed between said speed control and said second named valve whereby -the degree of pressure released by said 'second valve is regulated in part by the position of said speed control.

3. In combination, a throttle lcontrolled engine, a power transmission system embodying a uid pressure actuated clutch for establishing changes in speed ratio, a speed ratio control mechanism operatively connected with said system comprising, means constantly responsive to said throttle position, said means including mechanism jointly operative with a speed governor, .a control device effective upon said clutch subject to variations in the degree of engine vacuum, and means coincidentally operative with said engine throttle to vary the effective rate of engagement of said clutch.

4. In step-gear transmissions embodying gradually engageable friction devices, in combination withautomatic speed ratio changing mechanism, a movable control element including means responsive to variation in speed of the motor, means responsive to variations in torque demand upon said motor, and additional means responsive to variations in engine torque operative upon additionalk mechanism arranged to change thev rate of engagement of said devices during the ratio shift intervals established by said automatic speed ratio changing mechanism.

5. In automatic controls for motor driven vehicles, in combination, an engine, a driving shaft, a driven shaft, a variable speed transmission' arranged between the shafts for multiplying torque and embodying fluid pressure sustained clutch elements, an automatic control device therefor including a governor responsive to the speed of one of the shafts, operator controlled means connected to said device and operative to influence governor selection of speed ratio, said means coincidentally controlling said engine speed, and additional means controlled proportionally in accordance with movement of said first named means, operative to vary the rates of engagement of said clutch elements for predetermined torque operating conditions.

6. In automatic velocity controls for motor vehicles, in combination, a transmission speed ratio control element, speed responsive means continuously operative to influence the motion of said element, means varying with engine torque jointly operative to influence the motion of said element, and additional means controlled by said last named means, effective to modify the rate of change of established drive for predetermined power and load conditions.

7. In automatic velocity control systems for motor vehicles, a continuously operative speed ratio control element, fluid pressure actuated speed ratio changing mechanism connected to said element, speed responsive means operative to influence the motion of said element, means responsive to engine torque operative to influence the motion of said element, and additional operator controlled means arranged to modify the rate of change from one speed ratio to another according to predetermined conditions of load and speed.

8. In combination with a variable speed transmission mechanism embodying fluid pressure sustained clutch elements, a drive shaft and a driven shaft, means responsive to variations in speed of one of the shafts-and controlling the operation of said mechanism, engine vacuum actuated means for controlling the rate of engagement of at least one clutch element, and manually actuated means operative with said frst named means to modify the action of said speed responsive device.

9. In change speed transmissions for motor vehicles, a variable speed gearbox, a fluid pressure sustained clutch device capable of occupying an operative and a non-operative position, for different speed ratio settings of said gearbox, locking means for automatically engaging said device in operative position, manually controlled means for releasing said holding means, and means rell shaft, a variable speed to vary the degree of pressure sustaining saiddevice.

10. In variable speed mechanism for motor vehicles, in combination, an engine, a throttle control'for said engine, a driving shaft, a driven gearing arranged between said shafts including a friction clutch for coupling the shafts to rotate at the same speed, gear driving means for connecting the shafts to rotate at relative speeds, fluid pressure means for actuating the clutch simultaneously operative to release drive through said second named means, and valving subject to variations in engine vacuum in accordance with motion of said throttle effective to vary the clutch capacity according to torque demand as determined by the position of said throttle.

11. A plurality of variable speed ratio transmission units arranged to transmit drive between an engine and a load, a multiplicity of friction torque sustaining elements adapted to provide forward drive in said units at selected speed ratios, servo mean's operative to sustain drive by certain of said elements while withholding drive by certain others of said elements, a speed responsive device connected to Said Servo means, control means coacting with said device arranged to transfer drive from said rstnamed elements to said second named elements and reversely according to relative coacting forces between said means and said device, and an auxiliary vacuum responsive control effective to vary the torque capacity of said ilrst named elements by variations in-engine when said servo means is made selectively operative by said occasion.

12. A driving and a driven shaft, a friction clutch comprising two members arranged to establish drive therebetween, engagement and disengagement means for said clutch, control mechanism for said means, means arranged to estabf lish positivegeared drive between said shafts when said clutch is disengaged, responsive to changes in engine vacuum, and a device operative to proportion the torque capacity according to predetermined conditions of load and speed of driving said shaft.

13. In power transmission controls, in combination, variable speed gearing, a friction clutch associated with said gearing arranged rto -be actuated by fluid pressure, engagement rate control means for said clutch including a conoidal spring resistance, an. auxiliary yielding means subject to fluid pressure, and a valve controlled torque to regulate the rate of actuation by said fluid pressure.

14. In variable speed gearing control for automotive vehicles, in combination, a,v plurality of geared step-ratio transmission units arranged to connect a power'source and a load, a plurality of clutches adapted to change the speed ratios of drive of said its, iiuid pressure means eective to shift dri e thereby change speed ratio, and control mechanism for said means operative to vary the rate of transfer from one ratio to another according to predetermined condition of engine torque.

Y JOmi BOLZA.

to and from said clutches and 

